The 13th Scandinavian International Conference on Fluid Power, SICFP2013, June 3-5, 2013, Linköping, Sweden

Simulative Analysis and experimental Investigation of Common Rail Injection

Pumps lubricated with tailor-made Biofuels

S. Heitzig, S. Drumm and H. Murrenhoff

Institute for Fluid Power Drives and Controls (IFAS), RWTH Aachen University, Aachen, Germany

E-mail: Stefan.Heitzig@ifas.rwth-aachen.de,

Hubertus.Murrenhoff@ifas.rwth-aachen.de

Abstract

In the scope of the cluster of excellence “Tailor-Made Fuels from Biomass” new biofuels are

developed within an interdisciplinary research approach at RWTH Aachen University. The most

promising fuel candidates so far tend to have a low viscosity and a low lubricity compared to

diesel fuel. It is the task of the authors to develop a guideline for designing injection systems

that are especially adapted to these fuels. The main focus is on high pressure pump within

common rail systems. In this paper, the effects of those fuel candidates on the tribological

contacts in standard common rail pumps are investigated by means of experiments and

simulations. On the experimental side a test rig is presented, which allows the measurement of

the transversal force on a piston and the friction force, which occur in the piston/bushing-

contact. On the simulation side, elastohydrodynamic simulation models of several tribological

contacts within the standard pump are presented. Results of simulation and measurement are

presented.

Keywords: Elastohydrodynamic simulation, friction forces, tribological contacts, common rail

pump, fuels, journal bearings

1 Introduction and Motivation

1.1 Tailor-made Fuels from Biomass

Environmental pollution and the exploitation of non-

renewable resources of modern times force the development

of new sustainable biofuels. One approach to find such

biofuels is pursued within the cluster of excellence “Tailor-

Made Fuels from Biomass” at RWTH Aachen University.

Within the cluster, new biofuels are developed and tested in

an interdisciplinary approach, see fig. 1.

Figure 1: Tailor-made Fuels from Biomass

The fluids are evaluated regarding their suitability for use as

a biofuel on the basis of their conversion, combustion and

tribological characteristic. A lot of the investigated fuels

have an excellent combustion and canversion characteristic

but poor tribological properties. They can differ severely

from conventional fuels in their lubrication and viscosity

characteristics. Therefore, tribological problems and higher

efforts to pressurise the fuel to the injection pressure level

have to be expected in fuel-lubricated common rail pumps

when operated with the new fuel candidates. In order to

facilitate a widespread usage of potential new biofuels,

research at the Institute for Fluid Power Drives and

Controls (IFAS) focuses on two approaches to ensure a

proper functioning of the tribological systems. On the one

hand, low-lubricity biofuels can be blended with high-

lubricity biofuels in order to enhance their lubrication

characteristics in the boundary lubrication regime [1, 2]. On

the other hand, research on the components reveals what

extent modifications of those components (micro and macro

geometry, materials, coatings, etc.) can contribute reducing

the effort that is necessary in order to pressurise the fuel

candidates.

1.2 Common Rail Injection Pump

Diesel injection systems of modern passenger cars are

nowadays generally designed as so-called common rail

systems. In those systems the pressure build-up is decoupled

from the injection process by means of the so-called rail,

which functions as a hydraulic accumulator. In this paper

the focus is on the common rail pump, which is used to

pressurise the fuel to the injection pressure level. From the

437

hydraulic point of view this is a challenging task for several

reasons. Firstly, the pressure level of common rail systems is

much higher than in standard hydraulic applications.

Systems of the first generation started with a maximum

pressure level of 1350 bar. In the second and third

generation the maximum pressure level was increased in

order to achieve a more efficient combustion and fewer

emissions. To fulfill the upcoming regulation regarding

future emission limits, the next generation of common rail

systems will operate at even higher pressures from 2500 up

to 3000 bar. Secondly, from a tribological point of view the

fluid properties of diesel are worse in comparison to

standard hydraulic fluid. According to DIN EN 590, at

40 °C diesel has a viscosity within the range from 2,0 up to

4,5 mm2/s. This leads to more leakage losses in pumps and

more operation of the tribological contacts in the mixed and

boundary lubrication regime. And thirdly, those pumps are

for the rather cost sensitive automotive sector.

Figure 2: Common rail pump CP1. Source: Robert Bosch

GmbH

As a starting point the effects of the fuel candidates on a

standard common rail pump of the first generation are

studied in this paper. This kind of pump, a BOSCH CP1 [3],

is a radial piston type design and consist of three pistons, see

fig. 2. Via a metal sheet clamp, each piston is connected to a

piston slipper, which is sliding on a planar surface of the

polygon ring. This ring is connected via a rotational journal

bearing to the eccentric shoulder of the pump drive shaft.

All tribological contacts in this pump are lubricated by the

fuel itself.

Figure 3: Sequence of one revolution

In fig. 3 the kinematics of this kind of pump are illustrated

for the movement of one piston unit. The revolution starts at

0° drive shaft angel with the transition from the compression

stroke to the suction stroke. The piston is at the outer dead

center (ODC) and hence its velocity is zero. From 0° to 180°

the suction stroke is continuous and the piston moves

downwards. At 90° the direction of the relative motion in

the sliding contact between slipper and polygon changes its

direction. At 180° the piston is at the inner dead center

(IDC) of its movement. From 180° to 360° the compression

stroke is continuous and the piston moves upwards again.

At closer inspection of cross section of the pump in fig. 2 it

becomes obvious that there is a slight displacement d

between the axis of each piston and the axis of the drive

shaft. This detail in design, which can be found in every

common rail pump of this type, is exemplified in fig. 4.

Figure 4: Tilting of the polygon ring

The motivation for this displacement d is a reduction of the

tilting φ of the polygon ring. This tilting is determined by

the forces acting on the polygon ring. In the investigated

pump type each piston axis has a displacement d to the drive

axis of half the eccentricity e of the drive shaft. With this

displacement the tilting of the polygon ring is reduced and

hence the transversal load Fx on each piston, see eq. 1.

costan FFFFFzxxx

(1)

If no tilting occurs (φ = 0), the transversal load Fx is equal to

the friction force Fµ in the polygon/slipper-contact, because

the bearing surface on the polygon is orthogonal to the axis

of the piston. As soon as the polygon is tilted (φ ≠ 0), a

transversal force Fxφ due to the inclined plane has to be

taken into account.

1.3 Focus of this Paper

Within the scope of this paper simulative and experimental

approaches investigating the effects of the new fuel

candidates on standard common rail pumps are discussed.

The presented experimental approach tries to cover the

operating characteristic and conditions of standard pumps as

realistic as possible. But nevertheless, a few differences

occur between the operating characteristic and conditions in

the test rig and in a standard pump. This experimental

approach allows the measurement of the friction force in the

piston/bushing-contact as well as the transversal load on the

piston. In the simulative approach further contacts are taken

into consideration to reach a better comparability of

measurement and simulation. As this paper will

demonstrate, there are several specific reasons why a

satisfying comparability of measurement and simulation is

piston

drive shaft

slipper

polygon

bushing

SSsS suction stroke

SSsS compression stroke

S

S

s

S

0°/360° 90° 180° 270°

ODC

IDC

438

difficult to reach at this point. Improvements on the

experimental side will be exposed, which will enable a

better comparability in the future.

2 Experimental Approach

2.1 Test Rig Set-Up

Over the past decades several test rigs were designed to

measure the friction forces in tribological contacts within

hydraulic pumps [4] [5] [6] [7] [8]. The basic principle

chosen for this test rig is comparable to the one developed

by Breuer [9].

The design of the test rig is shown in fig. 5 (a). This design

allows the measurement of the friction force (z-direction)

and the transversal load (x-direction) of one piston unit. All

tribologically relevant parts are taken from standard pumps

(piston, bushing, piston slipper, polygon ring, drive shaft).

One bushing is separated from the housing and connected to

the force measurement platform. In order to compensate the

huge forces occurring because of the pressure build-up in

the piston chamber, a compensation piston with the same

diameter is installed. This piston transmits the pressure

forces directly to the base frame so that these forces are not

measured by the force sensors.

Figure 5: Test rig, layout (a) and hydraulic diagram (b)

Figure 5 (b) shows the hydraulic diagram of the rig. A

simple external gear pump is used as a fuel feeding pump.

To achieve a constant supply pressure of 6 bar a pressure

relieve valve is installed. Two check valves at each piston

are used for commutation. The flow of the pistons is

pumped to two different common rails (accumulators). One

rail is connected to the measured piston while the other rail

is supplied by the unmodified pistons. The pressure in each

common rail is controlled by a pressure relieve valve. By

means of the use of two common rails it is achieved that

only low pressure connections are installed between the

platform measuring the force and the base frame of the test

rig. That is how no additional forces develope because of

pressure pulsation in the high pressure system are induced to

the force measuring system.

A more detailed description of the test rig can be seen in

[10].

2.2 Comparison of Test Rig and Standard Pump

Even though operating conditions of the tribological

contacts in the test rig are similar to those in standard

pumps, there are a few differences. In the following, these

differences will be discussed.

In order to separate the pressure forces from the friction

forces a compensation piston is added to the test rig. This

will slow down the pressure build-up in the piston chamber,

because additional leakage occurs at this piston/bushing-

contact. In addition to that, friction forces will occur in this

contact as well. Even if there is no relative motion between

the surfaces in this contact, hydrodynamic friction forces

due to the leakage will occur at the bushing and will be

measured by the force sensors. In order to be able to install a

dynamic pressure sensor and the compensation piston, dead

volume is added to the piston chamber of the measured

piston unit. Like the additional leakage due to the

compensation this will slow down the pressure build-up as

well.

For the reasons explained above, two separated common

rails are used in the rig. Thus, the pressure level in the

measured piston/bushing unit is completely decoupled from

the pressure level in the other two piston units. The pressure

levels in the two rails are adjusted by manually adjustable

pressure relief valves. Slightly different pressure levels as

well as different dynamic characteristics will occur in the

two rails. Furthermore, in a standard pump the housing is

pressurised (roughly 3 – 6 bar) and flushed by the flow of

the feeding pump. At this point this is not the case in the rig.

In order to avoid a force shunt between the pump housing

and the force measurement platform, no sealing is installed

between the housing and the bushing. As a consequence it is

not possible to pressurise the housing of the rig. The

flushing of the housing in the rig is only achieved by the

leakage of the three piston units and not by the total flow of

the supply pump. A further difference between pump and

test rig is that the fuel in the rig is pressurised several times.

In a common rail system the pressurised fuel is burned after

compression. Only a small amount is pressurised several

times. In the rig an aging of the diesel can occur during the

measurement, because it is pressurised and throttled several

times.

Some of those differences will be eliminated in future

measurements, see Chapter 5 of this paper.

2.3 Measurement Results

In a first step, measurement with standard diesel were

conducted to prove the functioning of the rig and to detect

weaknesses. Measurement results at a selected point of

operation (rail pressure: 1800 bar; revolution: 200 1/min)

are shown in fig. 6.

Drive motor

Fly wheel

Angle sensor

Force sensor

Test box

Measuring platform

Compensation piston

Measured piston

Drive shaft

b) Hydraulic diagramm

Feeding pump

Common rail

a) Test rig layout

439

Figure 6: Measurement results

The revolution starts at the outer dead center of the piston

with a rotation angle of 0°. From 0° to 180° the piston is

moving downwards (suction stroke) and from 180° to 360°

upwards (compression stroke). Forces during the suction

stroke are on a low level due to the low level of the piston

pressure of just a few bars. Just at the very beginning of the

suction stroke there is still a higher piston pressure level and

hence a higher level of forces. This higher level of piston

pressure is due to the compressibility of the fuel and the

elasticity of the components. At roughly 80° the

decompression is completed and the piston pressure stays on

the pressure level of the feeding pump. As soon as the

compression stroke is entered, the pressure in the piston

chamber raises. Due to a great amount of leakage in the

piston/bushing-contact of the measured piston unit and the

compensating piston as well as due to the compressibility of

the fuel and the elasticity of the solid bodies, the raise of the

pressure takes some time. At roughly 270° the piston

pressure reaches the pressure level of the rail. From then on,

fuel is delivered to the rail. A few degrees before the suction

stroke is entered the pressure level in the piston chamber

decreases. At this point the piston movement is not fast

enough to compensate the leakage losses. The transversal

forces on the piston Fx are submitted to a change in direction

at 270°, because at that point the relative motion in the

polygon/slipper-contact changes. This is also the case at 90°,

but at that point the change of direction is not visible in the

plots due to the low level of the force.

As exemplified before, the transversal load on the piston is

not just the friction force in the polygon/slipper-contact. The

transversal load of the cylinder Fx is equal to that friction

force Fµ only if the surface of the polygon ring is orthogonal

to the piston axis, see eq. 1. To detect the influence a tilting

of the polygon may have on the transversal load,

measurements with different displacements of the axis of the

measured piston unit to the drive shaft axis were conducted.

The test rig allows the variation of the displacement of the

measured piston unit from -2,5 mm up to +2,5 mm. The

results of these measurements are shown in fig. 7 (rail

pressure: 1800 bar; revolution: 200 rpm/min).

Figure 7: Measurement results for Fx with different

displacements

As it can be concluded from the results the displacement has

a huge influence on the transversal load on the piston. The

displacement of -2,5 mm leads to the highest transversal

load on the piston. The transversal load is reduced with

every step the displacement takes towards 2,5 mm. It can be

assumed that the friction force Fxμ in the polygon/slipper-

contact is independent of the displacement and has

approximately the same characteristic in all five

measurements due to comparable tribological operating

conditions of this contact. Hence, the variation in transversal

load Fx on the piston is caused by a different tilting

characteristic of the polygon ring.

To draw conclusions from the measurement results

concerning the most benificial displacement d is not

possible due to several reasons. Firstly, in the rig the

displacement of only one piston unit is changed. The

position of the other two piston units cannot be changed.

Secondly, up to now only one operating point is measured,

which is not a typical operating point of this kind of pump.

It can be expected that the most beneficial displacement is a

function of the operating conditions (revolution, pressure,

temperature) as well as of the tribological properties of the

fuel (lubricity and viscosity characteristic).

3 Simulative Approach

In the simulative approach, elastohydrodynamic models of

the fuel-lubricated contacts were set up. In the following, a

short introduction to the simulation tool will be provided.

For more detailed information regarding the simulation tool

the reader is referred to Knoll et al. [11, 12].

Solid bodies are incorporated in the simulation by FEM-

models. The local deformation and global rigid motion is

determined by the integration of a special formulation of

Newton’s equation of motion, which combines large rigid

body motions and small deformations of the elastic bodies,

see eq. (2). The properties of the elasic bodies are

incorporated in this equitation by the mass (M), damping

(D) and stiffness (K) matrices, which are obtained from

FEM-models. Centrifugal, gyroscopic and coriolis forces

arising on the right hand side of eq. 2 besides external

No

rma

lise

d f

orc

e [

-]

Drive shaft angle [°] Drive shaft angle [°]

No

rma

lise

d f

orc

e [

-]

d = 1,5 mm

d = 0 mm

d = 2,5 mm

d = -1,5 mm

d = -2,5 mm

rail pressure

Fx

Fz

piston pressure

440

forces. The total acceleration vector ( ̈ ̈) is separated

into rigid body ( ̈) and elastic accelerations ( ̈).

t+t+t+t=+++gycoceext

ffffuKuDuqM (2)

Prior to the elastohydrodynamic simulation, a reduction

scheme is applied in order to achieve acceptable

computation time by limiting the number of degrees of

freedom (DOFs) of conventional finite element models.

With the help of the reduction step, the modeling

capabilities of the structure are reduced to reasonable

deformation modes. It is important to choose relevant DOFs

for the reduction step, because all solutions in the reduced

system are weighted by superposition of these modes. In this

reduction procedure, which is called mixed Guyan/modal

reduction scheme, two different kinds of DOFs have to be

chosen. The first group, the so-called static Ritz vectors

model local deformations in stiff regions. These DOFs are

used to cover the local deformations of for example bearing

surfaces or surfaces were a load is applied. The second kind

of DOFs are the so-called modal Ritz vectors. These DOFs

represent the eigenvectors of the structure, which are

obtained from a modal analysis. With the help of these

DOFs global deformations can be represented by the

reduced structures. The static and modal Ritz vectors

represent the reduced DOFs and called static and modal

DOFs, respectively. The reduction step is done once for

each structure before the elastohydrodynamic simulation.

Besides standard coupling of different solid bodies by

means of springs and dampers the solid bodies can be

coupled via nonlinear hydrodynamic reaction forces which

represent lubricating films. Those hydrodynamic forces are

calculated on a separate 2d-mesh by solving Reynolds

equation for thin lubricating films, see eq. 3.

Microhydrodynamic effects are incorporated via flow

factors according to Patir and Cheng [13], which are gained

from measurements of real surface segments.

3;

12

3

i=

x

ph

xj

p

ij

i

t

h

x

vv

x

hvv

j

s

ijVii

i

ii

22

1212

(3)

Cavitation is considered by Reynolds boundary condition.

The different fuels are characterised by their viscosity

characteristics. Since high pressures and temperatures can

occur in injection systems, the temperature and pressure

dependency of the viscosity plays a major role. In order to

determine the fuel viscosity characteristic, a viscometer for

elevated pressure and temperature has been set up at IFAS

[14]. Based on these viscosity measurements, models were

parameterised and included in the simulation, see eq. 4.

11

0

0

0

z

p

p

z

p

e

(4)

As soon as the hydrodynamic forces cannot bear the entire

external forces on the lubrication film, solid body contact

occurs. The contact pressure is determined with the help of

the contact pressure curve. This curve is calculated based on

measurements of real surface segments. If a critical gap

height, which depends on the surface roughness of the parts

sliding against each other, is reached during the dynamic

analysis, solid body contact occurs. Friction forces are

calculated according to eq. 5.

dA

z

udApF

cccfriction

(5)

Up to now an energy equitation is not included in the

simulation tool and hence local temperatures cannot be

calculated by the simulation tool.

4 Simulation Models

4.1 First Simulation Models

In a first step simple simulation models for the

polygon/slipper-contact as well as for the piston bushing

contact were set up. In order to reduce simulation time,

those two lubricating films were simulated in separate

models. The simulated friction force in sliding direction

(polygon/slipper simulation) is transferred to the piston in

the piston/bushing simulation as transversal load. A more

detailed description of the models can be taken from [15].

The viscosity characteristic of the fuels is incorporated

according to eq. 4. The parameters are set according to

measurement results from a high pressure viscometer, which

was set up at IFAS. The results of those measurements for

several fuels at T = 40°C are displayed in fig. 8. The

intensive pressure dependency of some fuels becomes

obvious.

Figure 8: Pressure dependency of the viscosity at 40°C [16]

Fig. 9 shows the temperature dependency of some TMFB

candidates at atmospheric pressure. In Common-Rail-

Systems temperatures up to 120 °C and pressures up to

2000 bar have to be expected [17]. Hence, the temperature

as well as the pressure dependency cannot be neglected in

simulations.

441

Figure 9: Temperature dependency of the viscosity at

atmospheric pressure [16]

For the following simulations the promising fuel candidates

dibutylether (DBE), 2-Methyltetrahydrofuran (2-MTHF)

and butyl levulinate (BL) were chosen.

In fig. 10 simulation results for the bearing forces of the

contact are pictured. At closer inspection of the results for

the polygon/slipper-contact (left side of fig. 10) a huge

influence of squeeze-effects can be identified, which

reduces or even avoids solid body contact at the beginning

of the compression stroke at 180°. Those squeeze-effects are

viscosity-dependent and hence less distinctive for the fuel

candidates like DBE, 2-MTHF and BL with a low viscosity.

As a consequence, more solid body contact occurs if this

contact is lubricated with those fuel candidates This leads to

higher friction forces in the polygon/slipper-contact and

consequently to a higher transversal load on the piston in the

piston/bushing simulation. There, because of this higher

load and due to the lower viscosity, more solid body contact

occurs in this contact as well. All in all these results indicate

that this kind of pump design is sensitive to fuel viscosity.

Figure 10: Test rig, layout (a) and hydraulic diagram (b)

The comparability of those models to a real pump is limited.

For example the tilting of the polygon ring is not covered by

the models and hence the transversal load on the piston is

just the friction force Fxµ of the polygon/slipper-contact.

Furthermore, an almost ideal shape of the pressure in the

piston chamber was assumed. From 5° to 175° a constant

suction pressure level of 3 bar was assumed and from 185°

to 355° a constant injection pressure level. From 355° to 5°

and from 175° to 185° the commutations in-between the

pressure levels were assumed. As measurements at the test

rig have shown, the pressure in the chamber raises slowly

due to leakage, compressibility and elasticity.

4.2 Improved Models

In a next step those models were extended to achieve a

better comparability of the measurement and the simulation.

In the following these models – all in all three of them - are

presented. In a first simulation the tilting φ of the polygon is

detected. This tilting φ is an input for the second simulation.

In this simulation the contact between polygon and slipper is

modelled more accurately. The output of that simulation is

the transversal load Fx on the piston. The piston bushing

contact is calculated in the last simulation. The reason for

three separate models instead of one model for the whole

pump is once again the required computation time.

4.2.1 Tilting of the Polygon Ring

As measurements indicated, the tilting φ of the polygon ring

has a huge influence on the transversal load Fx on the piston,

see fig. 3. To consider this in the simulation, an additional

elastohydrodynamic simulation tool was set up, see fig. 11.

The piston pressure input, which is displayed in fig. 11 as

well, was determined a priori in a separate hydraulic

simulation with the simulation tool DSHplus. This model

consists of eight rigid or non-rigid solid bodies:

piston (three times); rigid (0 DOFs)

slipper (three times); non-rigid (50 DOFs)

polygon ring (one time); non-rigid (100 DOFs)

eccentric shoulder (one time); non-rigid (50 DOFs)

Figure 11: Simulation model for the tilting of the polygon

The eccentric shoulder is coupled with the polygon via a

rotational journal bearing (140 nodes) and the three slippers

piston

slipper

ω

p1(shaft angle)

p2(shaft angle)

crank shaft angle [deg]

p3(shaft angle)

polygon

eccentric

shoulder

pis

ton p

ress

ure

[bar

]

442

are coupled to the polygon via axial journal bearings (71

nodes). All bearings are lubricated with fuel. The way the

piston is coupled to the slippers has a high influence on the

tilting of the polygon. In the pump, the piston is held down

to the slipper by a small metal clip, see fig. 12. This clip

does not block a tilting up to a tilting angle of 4° in both

directions. As soon as piston and slipper axis are not

concentric, not the whole bottom surface transfers the load

from the piston to the slipper. On one side of the contact a

gap arises and on the other side the load is transferred via

solid body contact pressure of the contacting surfaces. In the

simulation this characteristic could be modelled with the

help of a lubricating film. But since the computation effort

to solve the Reynolds equitation on this film and to

determine the contact pressures on the bearing surfaces is

huge, this option is not implemented in the simulation

model. Instead another solution, which reduces computation

time dramatically, is chosen. Each piston is coupled with its

slipper via three compression springs, see fig. 12. With this

kind of coupling the basic characteristic of the contact

between piston and slipper in the pump can be modelled,

assuming an adequate parameterisation of those three

springs.

Figure 12: Coupling of piston and slipper

4.2.2 Polygon/Slipper-Contact

The second simulation is a more precise simulation of the

polygon/slipper-contact, see fig. 13. In this simulation only

one lubricating fuel film is covered and there are only three

solid bodies:

piston; non-rigid (0 DOFs)

slipper; non-rigid (250 DOFs)

1/3 polygon; non-rigid (300 DOFs)

Figure 13: Simulation model for the polygon/slipper-contact

The polygon is tilted as a function of the shaft angle

according to the results of the first simulation. In order to

get more accurate results the number of DOFs of the slipper

and polygon is increased in comparison to the first

simulation. Furthermore, the 2d-mesh representing the

lubricating film between slipper and polygon is defined

more precisely (1871 nodes). The piston is modeled as a

rigid body. The coupling between piston and slipper is

implemented by three compression springs again. This

simulation delivers as a result the transversal load Fx on the

piston as a function of the shaft angle.

4.2.3 Piston/Bushing-Contact

The third and last model is a more accurate model of the

piston/bushing-contact, see fig. 14.

Figure 14: Simulation model for the piston/bushing-contact

The solid bodies, the piston and the bushing, are modeled as

non-rigid bodies with 100 DOFs and 300 DOFs

respectively. The lubricating film along the piston/bushing-

contact is represented by a 2d-mesh with 1936 nodes. Along

the gap in axial direction the pressure level is reduced by

throttling effects. At the top end of the piston the pressure is

on a high level during the compression stroke and at the

lower end the pressure level is equal to the pressure level in

the pump housing. The fuel is throttled from the piston

chamber pressure level to the pressure level in the housing,

as it can be seen in the fuel pressure distribution in fig. 13.

In real pumps this energy dissipation leads to a heating of

the fuel along the gap. Since the energy equation is not

implemented in the model, thermal effects are included in a

simplified manner. As measurements in literature show [18],

the temperature distribution along the bushing is

approximately linear. This knowledge is used to define a

temperature-dependent reference viscosity η0 along the z-

direction of the lubricating gap.

4.3 Simulation Results

In the following some results of the improved models are

presented. The simulated operating point is the same like in

fig. 10 (rail pressure: 1350 bar; revolution: 2500 1/min,

temperature: 50 – 100°C).

bushing

piston

slipper

φ

φ

0L

EAc i

i

0ic

p1(shaft angle)

x0

compression springs

clamp

p1(shaft angle)

p1(shaft angle)

piston (rigid)

slipper (non rigid)

pressure distribution

1/3 polygon

(non rigid) φ (shaft angle)

piston (non rigid)

fuel pressure

contact pressure

bushing (non rigid)

443

Figure 15: Results regarding the tilting of the polygon

In fig. 15 the simulated tilting of the polygon is plotted. A

periodic characteristic is visible. Every 120° a peak up to

1.8° occurs due to the commutation. The dashed line

represents the simulation results for the tilting of the

polygon in the test rig. This tilting differs from the tilting in

a standard pump due to the deviating pressure build-up in

the measured piston unit (piston 1) in comparison to the

other two piston units in that test rig. As explained before,

the pressure build-up in the measured piston unit is slowed

down by leakage at the compensation piston and by an

increased dead volume.

Figure 16: Results regarding friction forces of the

polygon/slipper-contact and the transversal load on the

piston

Figure 16 shows the friction forces in the polygon/slipper-

contact and the transversal load Fx on the piston. Besides the

total friction force Fµ its share resulting from solid body

friction and its share due to hydrodynamic friction are

plotted. The hydrodynamic share is neglectable so that

almost all the friction in this contact is due to solid body

friction. Because a tilting of the polygon occurs, the

transversal load deviates from the total friction force. Only

when the tilting angle is zero the transversal force Fx is

identical to the friction force Fµ, e.g. at 325°.

Comparing the simulation results for the transversal load on

the piston of the first simulation models plotteted in fig. 10

(upper left side) to the results of the improved models it is

visible that the transversal load simulated with the improved

models is oscillating due to tilting of the ring. This

characteristic is not visible in the results of the first models

in fig. 6. There the transversal load is steadily rising during

the compression stroke. In the measurement results plotted

in fig. 6 and 7 an oscillating characteristic similar to the

results of the improved models is visible, too. Neverthless, a

comparison of measurement and simulation is disputable at

this point because of different operating points in the

simulation and the measurement.

Figure 17: Results regarding contact forces of the polygon

slipper-contact

In fig. 17 the contact forces in the polygon/slipper-contact

are presented. Again distinctive squeeze-effects at the

beginning of the compression stroke are visible. Solid body

contact is avoided or reduced by these effects at the

beginning and hence the friction force in x-direction starts

on a low level. Later on, this effect decreases and

consequently more and more of the load is carried by solid

body contact.

Figure 18: Results regarding friction forces of the

slipper/polygon-contact

Figure 18 presents the friction forces in the piston/bushing-

contact. Again, the total friction force is divided into its

hydrodynamic friction share and its contact friction share. In

contrast to the polygon/slipper-contact, the hydrodynamic

444

friction share is not negligible. Because of the huge pressure

difference during the compression stroke of 1350 bar an

intensive poiseuille flow leads to a hydrodynamic friction

force of roughly 20 N. Because of the transversal load on

the piston solid body contact occurs and hence a solid body

friction force.

5 Conclusion and Outlook

In this paper an experimental and simulative approach for

investigating the impact of new biofuels on the tribological

contacts in common rail pumps was presented. Potential for

further improvements on the experimental side and first

results were discussed. These results are indicating a huge

influence of the tilting φ of the polygon on the transversal

load Fx on the piston. On the simulative side first simulation

models and results for different fuels were presented. Those

first models do not capture the tilting of the ring. In order to

achieve a better compatibility with the measurements, the

simulation models were improved and extended. The results

for diesel show a better compatibility to the measurements,

but a validation of the simulation by the measurement is not

possible at this time, because there are several drawbacks

mainly on the experimental side. To further increase

comparability, improvements of the test rig will be done.

This includes improvements of the pressure control in the

rail, a better capturing of local temperatures and leakage

measurements at the piston or at the compensation piston.

After those modifications, testing of more typical operating

points at higher speeds of more than 1000 rev/min will be

possible as well as testing of different fuels.

6 Acknowledgement

This work was supported by the Deutsche Forschungs-

gemeinschaft (German Research Foundation) within the

cluster of excellence 236 “Tailor-made Fuels from Biomass”

[EXC 236].

Nomenclature

Designation Denotation Unit

A Area [m2]

fce Centrifugal force [N]

x Coordinate [m]

fco Coriolis force [N]

D Damping matrix

ρ Density [kg/m3]

d Displacement [m]

e Eccentricity [m]

E E-module [N/m2]

fext External force [N]

Φ Flow factor [-]

Fi Force [N]

µ Friction coefficient [-]

h Gap height [m]

fgy Gyroscopic force [N]

L Length [m]

M Mass matrix

p Pressure [bar]

c Spring stiffness [N/m]

K Stiffness matrix

T Temperature [°C]

φ Tilting angle [°]

t Time [s]

v Velocity [m/s]

η Viscosity [(N s)/m2]

References

[1] A Fatemi, K Mausch, H Murrenhoff, K Leonhard.

Experimental investigation and theoretical prediction of

the lubricity of biofuel components. In: Fluid power and

motion control: (FPMC 2010). Basildon : Hadleys,

2010. – ISBN 978–0–86197–181–4, pp. 39–51

[2] K Mausch, A Fatemi, H Murrenhoff, K Leonhard. A

COSMO-RS based QSPR model for the lubricity of

biodiesel and petro-diesel components. In: Lubrication

Science 23 (2011), Nr. 6, pp. 249–262. – ISSN 1557–

6833

[3] K Reif. Moderne Diesel-Einspritzsysteme. Wiesbaden :

Vieweg + Teubner Verlag / GWV Fachverlage GmbH,

2010. – ISBN 978–3834813121

[4] K-Th Renius. Untersuchungen zur Reibung zwischen

Kolben und Zylinder bei Schrägscheiben-

Axialkolbenmaschinen. In: VDI Forschungsheft 561,

VDI-Verlag, Düsseldorf, 1974

[5] O Koehler. Anlaufreibungsverluste eines

Schrägscheiben-Axialkolben-Motors. In: Ölhydraulik

und Pneumatik Nr. 31, Jahrgang 11, 1987

[6] S Donders. Kolbenmaschinen für HFA-Flüssigkeiten –

Verlustanteile einer Schrägscheibeneinheit. Verlag

Mainz, Aachen, 1998

[7] A Kleist. Berechnung von Dicht- und Lagerfugen in

hydrostatischen Maschinen. Shaker Verlag, Aachen,

2002

[8] R Lasaar. Eine Untersuchung zur mikro- und

makrogeometrischen Gestaltung der Kolben-

/Zylinderbaugruppe von Schrägscheibenmaschinen.

VDI Verlag, Düsseldorf, 2003

445

[9] D Breuer. Reibung am Arbeitskolben von

Schrägscheibenmaschinen im Langsamlauf. Shaker

Verlag, Aachen, 2007

[10] S Drumm, S Heitizg, H Murrenhoff. Friction

measurement of plunger bushing contact in fuel

injection pumps. Proc. of Bath/ASME Symposium on

Fluid Power & Motion Control, ISBN: 978-0-86197-

187-9, S./Art.: 51-58

[11] G Knoll, R Schönen, K Wilhelm. Full Dynamic

Analysis of Crank-shaft and Engine Block with Special

Respect to Elastohydrodynamic Bearing Coupling. In:

Proceedings of the ASME Internal Combustion Engine

(ICE) Division Meeting Bd. 28, 1997, pp. 1–8

[12] R Schönen. Strukturdynamische Mehrkörpersimulation

des Verbrennungsmotors mit elastohydrodynamischer

Grundlagerkopplung. Kassel: Kassel Univ. Press, 2003

(Fortschrittsberichte Strukturanalyse und Tribologie,

Teil 15). – ISBN 3–89958–507–0

[13] N Patir, H S Cheng. Application of Average Flow

Model to Lubrication Between Rough Sliding Surfaces.

In: Journal of Lubrication Technology 101 (1979), Nr.

2, pp. 220–229

[14] S Drumm, A Wohlers, A Fatemi, H Murrenhoff.

Viscosity and bulk modulus measurements under high

pressure conditions. In: Society of Tribologists and

Lubrication Engineers annual meeting and exhibition

2010: Las Vegas, Nevada, USA, 16 - 20 May 2010. Red

Hook, NY : Curran, 2010. – ISBN 978–1–61738–727–

2, pp. 413–416

[15] S Heitzig, S Drumm, L Leonhard, H Murrenhoff, J

Lang, G Knoll. Simulative Analysis of the Tribological

Contacts of Common Rail Injection Pumps Lubricated

by Tailor-Made Biofuels. In: Tribologie und

Schmierungstechnik

[16] S. M. Drumm. Entwicklung von Messmethoden

hydraulischer Kraftstoffeigenschaften unter Hochdruck.

Shaker Verlag, Aachen, 2012

[17] T. Illner,. Oszillierendes kippbewegliches

Axialgleitlager bei Grenzreibung und

Kraftstoffschmierung. Shaker Verlag, Aachen, 2010

[18] J Blume. Druck- und Temperatureinfluß auf Viskosität

und Kompressibilität von flüssigen Schmierstoffen,

RWTH Aachen, Diss., 1987

446